This invention relates to rotating shafts, of the kind that have a tendency to undergo whip-type vibrations, in service. Such tendencies can occur when, for example, the distance between the shaft bearings is long, and there is a heavy mass on the shaft. In these cases, the critical rotational speed of the shaft, at which the shaft can start to undergo high-amplitude vibratory excursions, is low enough to be within the range of operating speeds.
In many rotating-shaft machines, the critical speed of a shaft is far higher than the speeds that can possibly be reached in service, and so no precautions need be taken to protect or contain the shaft against high-amplitude vibrational excursions. But when the shaft is likely to approach its critical speed during service, the designer must consider taking precautions.
These precautions can take the form of a shaft confinement system, by which the shaft is mechanically constrained against moving more than a predetermined transverse or lateral or radial distance. The shaft confinement system might comprise constrainment-rings, for example, or the like, which surround the shaft, and the designer must of course see to it that the rings can cope with the rotating shaft bouncing against, and rubbing against, the constrainment-rings.
The high-amplitude vibratory excursions are measured radially relative to the rotational axis of the shaft. The mode of vibration giving rise to the excursions is often simple bending of the shaft, i.e bending of the shaft between supports defined by the shaft bearings. But in other cases the mode of vibration giving rise to the excursions may be, for example, torsional. The critical speed in question is usually defined by the fundamental frequency of the vibrations, but it may be a harmonic. In many cases, the critical speed is determined by the simple-bending fundamental, but once the shaft starts to bounce against the confinement system, failure may occur due to the resulting high torsional stresses induced in the shaft.
In the case of a Francis turbine, for example, the geometry of the water-duct layout associated with the turbine often means that the shaft, which carries the (heavy) turbine rotor, has to be mounted between bearings which are spaced very far apart. Designers know that the precautions taken, in many Francis turbine installations, against critical-speed vibrations are not enough, and the installation may be troubled by packing-seals blowing out, bearings failing, and by the turbine rotor striking against its casing. Often, the amplitude of the vibrations can be reduced by reducing the power output of the turbine, but of course that is not a satisfactory solution. Often, the speed at which the turbine runs cannot be changed, being set by the electricity generating requirements of the installation.
The need for a shaft confinement system arises when the critical speed of a rotor is low enough to be within the service range of speeds at which the shaft will operate. While this can occur in other kinds of machines, the shaft confinement system of the invention will be described as it relates to a Francis turbine.
In the case of a double-ended Francis turbine, for example, the fundamental vibration of the shaft, in simple bending, might easily have a frequency as low as 10 Hz, whereby the critical speed of the shaft is nominally around 600 rpm. That is well within the range at which it is desired to operate the turbine, so the possibility of vibration problems is very real.
The designer can always protect a shaft from any tendency to whip-type vibrations, by making the critical speed of the shaft far above the operating speed, e.g by adding more bearings, making the shaft thicker, etc. But sometimes, the bearings have to be far apart, and the shaft has to be slender, and the shaft has to carry a heavy weight at the midpoint between the bearings, whereby the critical speed is low enough that the speed at which the shaft is operated is uncomfortably close to the critical speed. The invention is aimed at being useful in those cases. The invention is aimed at limiting the amplitude of whip-type vibrations, and thus at permitting a shaft to rotate at or close to its critical speed.
In the context of the invention, it is important to distinguish between the bearings that guide and support the rotating shaft, from the structures that confine the shaft and prevent it from undergoing excessive vibrations. The shaft confinement structure is characterized as a structure that does not support the shaft, and indeed this structure does not even touch the shaft, unless the shaft is undergoing radial excursions. The shaft-bearings, on the other hand, support and guide the shaft at all times, to the extent that it is the bearings that define the axis of rotation of the shaft.
In the invention, the shaft confinement structure includes a series of many tappets, which are pitched around the shaft, or rather, around a rotor component mounted on the shaft. The tappets are mounted in a fixed housing component surrounding the rotor, and protrude inwards, for operative engagement with an outwards-facing surface of the rotor. The tappets are adjustable, and are spaced a small distance away from the outwards-facing surface of the rotor. The tappet-clearance is sufficient that the inevitable (small) run-out of the rotor does not cause the outwards-facing surface of the rotor to touch the tappets, whereby the rotor normally runs clear of the tappets, but the tappet-clearance is small enough that the tappets serve to catch the shaft if the shaft should undergo a radial excursion, before vibrations can build up to a damaging amplitude.
With the invention, the designer need not be so conservative about designing the installation so that the installation is not troubled by approaching or reaching the critical speed. In many cases, the designer can even afford to deliberately include the critical speed within the service range.
The material of the tappet is important. A plastic material is preferred, for its shock-absorptive qualities. Metal is not preferred as a material for the tappets; the shaft will bounce off even a bearing metal, such as bronze, with much of its energy of vibration still present, whereas a plastic tappet tends to absorb and dissipate the energy of the impact. The energy is transformed into heat, of course, but the tappet does not become overheated, in a water turbine, as the water carries away the heat.
The plastic tappets should be dimensioned such that the wear rate of the tappet would be quite high, if just one tappet were to receive the impact of the bouncing rotor. Enough tappets should be provided, around the circumference of the rotor, that the impacts are spread over three or four tappets. In fact, it is preferred that the tappets be dimensioned such that, if one tappet should take the whole impact, that that tappet should wear rapidly; thus, if one tappet should be over-adjusted, it will be the first to touch the rotor, and will quickly wear away, until adjacent tappets start to touch the rotor.
It is preferred, not that the plastic material of the tappet should be springy or resilient, but, as mentioned, that the shaft should bounce back, off the tappet material, with much of its energy dissipated. As a result, the wear rate of a shock-absorptive plastic material, such as polyethylene, as actually measured, can be considerably less, in this mode of use, than the wear rate as measured when using a solid metal material. The rubbing speed of the tappet against the rotor can be very large, e.g. around twenty-five meters/sec, whereby the wear rate of the tappet can be very rapid unless the tappet material has some xe2x80x9cgivexe2x80x9d to it, which limits the impact force.
Also, the material should not be so ductile that the contact-tip of the tappet might tend to mushroom, or to bend over; rather, the effect of the rotor striking the tappet preferably should be for the material of the tappet to simply wear away, by abrasion, without the material distorting in any way. While polyethylene is a satisfactory material, of course, materials other than polyethylene should not be ruled out, in this context.
The size of the contact-tip of the (plastic) tappet is important, since the wear rate of the tappet is affected by the area of the tip. The area of the contact tip should be co-ordinated with the characteristics of the material, to produce a wear rate of the tappet in which one single tappet taking impacts from the rotor on its own is rapidly worn down, but once the rotor strikes three or four tappets, together, the wear rate becomes much smaller.
The tappets should be adjustable as to the clearance between the contact tip and the outwards-facing surface of the rotor. Suitable adjustment mechanisms are described herein.
By way of further explanation of the invention, exemplary embodiments of the invention will now be described with reference to the accompanying drawings, in which:
FIG. 1 is cross-sectioned side elevation of a Francis turbine, having a shaft confinement system that embodies the invention (the system is not visible in FIG. 1);
FIG. 2 is a close-up of the shaft confinement system of the Francis turbine of FIG. 1;
FIG. 3 is a correspondiong close-up of another shaft confinement system that embodies the invention;
FIG. 4 is a cross-section of a single-ended Francis turbine, having a shaft confinement system that embodies the invention;
FIG. 4a is a close-up of a portion of FIG. 4.